Orbital transmission with geared overdrive

ABSTRACT

The transmission includes an orbital gear complex in combination with a variable hydraulic pump and motor. The input to the transmission is increased in speed by the orbital gearing such that, when the pump and motor are not operating, the orbiter is stationary, and the orbital gearing produces an overdrive condition. A gear reduction is accomplished by rotating the web with the web-rotating device, providing a high gear reduction. The pump and motor are preferably long-piston hydraulic machines with infinitely variable swash plates. The hydraulic machines preferably have wobblers stabilized by full gimbals and hold-down plates with elongated holes for the long pistons to eliminate possible impacts between the hold-down plates and the head ends of the long pistons when the swash-plates are at or near their maximum angle of inclination.

REFERENCE TO RELATED APPLICATIONS

The subject matter of this application is related to the subject matterin co-pending U.S. patent application Ser. No. 10/789,739, entitled“LONG-PISTON HYDRAULIC MACHINES”, filed Feb. 27, 2004. This applicationis herein incorporated by reference.

The subject matter of this application is also related to the subjectmatter in co-pending application entitled “DUAL HYDRAULIC MACHINETRANSMISSION”, filed on the same day as the present application. Thisapplication is herein incorporated by reference.

BACKGROUND OF THE INVENTION

1. Field of the Invention

The invention pertains to the field of automotive transmissions. Moreparticularly, the invention pertains to an automotive transmission withorbital gearing and a variable web-rotating device.

2. Description of Related Art

Hydraulic pumps and motors with adjustable swash plates have beendiscussed for use in automotive transmissions for decades, but therehave been difficulties in building a hydraulic pump and motor that isboth lightweight and powerful enough at the speeds and pressuresnecessary for use in an automobile. Traditionally, hydraulic pump/motorswith adjustable swash plates have a fixed swash and a rotating cylinderblock. This arrangement works well for pump/motors for applications suchas golf carts and machinery, but for high pressure, high speed use inautomobiles a rotating cylinder block is too large, too heavy, and tooinefficient. While prior art relating to hydraulic machines hasdisclosed stationary cylinder blocks and split swash plates for nearly acentury, no designs have proven commercially successful for use with thecombination of high speeds and pressures required for automotive drives.The problem has primarily resided in the difficulty of providing asufficiently stable piston/swash interface.

In U.S. Pat. No. 5,440,878, “VARIABLE HYDRAULIC MACHINE”, issued Aug.15, 1995 to Gleasman et al. the piston-swash interface problem isaddressed. Long dog bones interconnect the pistons and the swash, and toprevent the collapse of the dog bones under rotational stresses, thewobbler of the swash is supported by a gimbal structure. A full gimbalstructure may be used with the fixed angle swash plates used on themotors, but a half-gimbal is used on the variable angle pumps.Prototypes of these machines exhibited undesirable vibrations andpulsations, indicating that the hydraulic machine could be improved.

In U.S. Pat. No. 5,513,553, “HYDRAULIC MACHINE WITH GEAR-MOUNTEDSWASH-PLATE”, issued May 7, 1996 to Gleasman et al., an alternative tothe gimbal is described. A spherical gear with spherical gear teethmeshes with gear teeth on the wobbler to stabilize the dog bones andwobbler by providing additional points of contact in comparison to thehalf-gimbal. This design, however, proved to be complicated tomanufacture.

In U.S. Published Application No. 2004/0168567, “LONG-PISTON HYDRAULICMACHINES”, published Sep. 2, 2004 to Gleasman et al., the dog bones arereplaced with long pistons and the gimbal and spherical gearing areeliminated. Spring pressure “hold down” is used to maintain the shoes ofthe long pistons in contact with the wobbler. No restraint is requiredto prevent collapse, since the long pistons do not collapse underrotational stresses or in the absence of hydraulic pressure. However,there is rotational stress placed on the wobbler by the high speed ofrotation of the rotor, and the effects of this stress cause undesirableinertial rotation of the wobbler.

There is a need in the art for variable hydraulic pumps and motorspowerful, efficient, lightweight, and small enough to be appropriate forautomotive transmission use.

In U.S. Pat. No. 6,748,817, “TRANSMISSION WITH MINIMAL ORBITER”, issuedJun. 15, 2004 to Gleasman et al., a variable pump and motor are combinedwith a gear orbiter to form an infinitely variable transmission. In thistransmission, as the speed of the hydraulic motor increases, the outputshaft speed increases and the speed of the vehicle increases.

Although an internal combustion engine is the industry standard forautomobiles in the United States, several major automobile manufacturersare researching a homogeneous-charge-compression-ignition (HCCI) engine.In a conventional gasoline engine, the air-fuel mixture is ignited by aspark plug to create power. In an HCCI engine, similar to in a dieselengine, a piston compresses the air-fuel mixture to increase itstemperature until it ignites. It is estimated that an HCCI engine iscapable of a 30% increase in fuel economy over a standard gasolineinternal combustion engine. A major hurdle for implementation of HCCItechnology in automobiles is a difficulty in controlling the combustionat low and high engine speeds.

There is a need in the art for a transmission, which provides thenecessary power to run an automobile while allowing its engine speed toremain in a relatively narrow low-to-moderate range where the combustionin HCCI engines is more easily controlled. Such a transmission allowsimplementation of more fuel efficient HCCI engines on gasoline-poweredvehicles.

SUMMARY OF THE INVENTION

The transmission includes an orbital gear complex in combination with avariable hydraulic pump and motor that operates unlike known automotivetransmissions. Contrary to conventional transmissions: when the webrotating device rotates the web at its highest speed in the samedirection as the engine, the transmission produces a reverse output atits highest speed; then, when the web is moved at a slightly slowerspeed, the output of the transmission produces neutral (no output);thereafter, as the rotation of the web in the direction of the engine isfurther slowed, the transmission produces a continuously decreasing gearreduction. When the web is at rest, the transmission provides anoverdrive condition. When the web is rotated in a direction opposite tothe engine, the transmission provides continuously higher overdriveratios. The pump and motor are preferably long-piston hydraulic machineswith infinitely variable swash plates. The hydraulic machines preferablyhave wobblers stabilized by full gimbals and hold-down plates withelongated holes for the long pistons.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 shows a partially schematic and cross-sectional view of ahydraulic machine with a variable swash plate angle. This hydraulicmachine is used as both the preferred pump and as the preferred motorfor this invention.

FIG. 2 shows a partially schematic and cross-sectional view of thehydraulic machine of FIG. 1 taken along the plane 2-2 with parts beingomitted for clarity.

FIG. 3A shows a partially schematic view of a hold-down plate, when theswash plate is inclined at +25°, as seen from the plane 3A-3A of FIG. 1.

FIG. 3B shows a partially cross-sectional view of the swash plate andpiston hold-down assembly, the view being taken in the plane 3B-3B ofFIG. 3A.

FIG. 4 shows a cross-sectional view of a single cylinder with a longspring.

FIG. 5 shows a partially schematic and cross-sectional view of ahydraulic machine with a split swash plate.

FIG. 6 shows a view of a “closed loop” arrangement of two hydraulicmachines as known in the prior art.

FIG. 7 shows a hydraulic machine of the present invention with a fullgimbal.

FIG. 8 shows a hold-down plate of the present invention with elongatedholes, the view being taken along plane 8-8 of FIG. 7.

FIG. 9 shows an orbital transmission in an embodiment of the presentinvention.

FIG. 10A is a schematic, approximately to scale, layout of theinvention's modular hydro-mechanical transmission in place behind astandard automotive engine.

FIG. 10B is an end view of the modular transmission shown in FIG. 10A.

DETAILED DESCRIPTION OF THE INVENTION

A transmission according to the invention includes a gear complex withan orbital web in combination with a web-rotating device for varyinggear ratios. Preferably, the web-rotating device is a variable hydraulicpump and motor. The input to the transmission is increased in speed bythe orbital gearing such that, when the pump and motor are notoperating, and the orbiter is stationary, the orbital gearing producesan overdrive condition. A gear reduction is accomplished by rotating theweb with the web-rotating device, providing a high gear reduction.

The transmission is appropriate for automotive use. While an orbitalgear complex of the present invention may appear to be similar to thecomplex disclosed in U.S. Pat. No. 6,748,817, the differences providesubstantially different results. The relative size of the input gear andits mating cluster gear are reversed. Instead of a conventional inputspeed reduction, the input speed is increased by the orbital gearing.Reduction of the input speed is controlled by the hydraulics, andoverdrive is achieved purely with the orbital gearing. This changeeliminates the need for an additional gear reduction and simplifies theoverdrive structure. In other words, when the hydraulic motor is stoppedbecause the hydraulic pump is at “zero” swash angle, the output shaft ofthe transmission is rotating faster than the input shaft.

The continuous and infinite-progression gear ratio change of theinvention's transmission occurs without any significant change in thespeed of vehicle's engine, and this continuous and infinite progressionextends through a remarkably wide range from a predetermined low gearratio (e.g., 22:1) up through an extended overdrive (e.g., as high as0.62:1) The engine may be maintained at a relatively low and efficientoperational level (e.g., 500 RPM) throughout the entire accelerationfrom a standing stop up to overdrive. This feature not only results infuel savings but, more importantly, in significant reduction ofpollutants. This is particularly true for diesel engine vehicles, sincethe engine's selected operational speed can be predetermined at a “sweetspot” which optimizes performance. As is well known, when a dieselengine operates at a constant speed, it discharges little, if any,pollutants.

This same feature can provide significant fuel savings despite the factthat the engine runs well below the “sweet spot” for which transmissionsof the prior art are designed. The sweet spot of an engine is theconventional optimal efficiency engine speed, which is the region of theefficiency map where the engine is most efficient at converting fuelinto mechanical power. For most automobile engines, the sweet spot isfound in the region of 1500 RPM (the region of maximum efficiency of thetorque converters of most conventional automatic transmissions). Atransmission of the present invention improves fuel economy despite thefact that it may maintain the engine at a speed where the engine is notrunning at optimum efficiency. This loss in efficiency is more thancompensated by the reduced fuel requirements of running at lower speeds(e.g., at 500 RPM rather than 1500 RPM). A further advantage of runningat such low engine speeds is a reduced pressure demand on the hydraulicpump and motor, thereby reducing the duty cycle on the hydraulics andfurther improving their durability.

In terms of fuel economy, an orbital transmission of the presentinvention brings city driving efficiency up to highway drivingefficiency. Since city driving accounts for about 60% of all driving,incorporation of a transmission of the present invention intoautomobiles should provide a significant fuel savings. Of course, sincetoday's engines are designed to run most efficiently in the range of1500 RPM, even further fuel savings may be achieved by combining atransmission of the present invention with an automobile engine designedto run most efficiently around 500 RPM.

A transmission of the present invention is capable of varying the speedof the drive shaft with minimal changes to engine speed. Thus, thepresent invention allows engine speed to remain in a relatively narrowlow-to-moderate range where the combustion in the recently proposed HCCIengines is more easily controlled. A transmission of the presentinvention is highly compatible with implementation of more fuelefficient HCCI engines on gasoline-powered vehicles.

The pump and motor of the invention are preferably long-piston hydraulicmachines with infinitely variable swash plates. Both hydraulic machinespreferably have split swash-plates that include a rotating and nutating“rotor” that is driven by the input axle and a nutating-only “wobbler”that rides on the surface of the rotor on bearings. In one embodiment ofthe invention, the bearings are needle bearings. The sliding shoes onthe long piston heads move in a “figure eight” path over the surface ofthe nutating wobbler. However, during the relative motion of the slidingshoes on the wobbler, this “figure eight” is really a lemniscate inthree dimensions on the surface of an imaginary sphere having a diameterthat decreases as the angle of the swash increases.

In long-piston hydraulic machines of the present invention, the wobbleris stabilized by a full gimbal. As the wobbler nutates, the distributionof piston shoe pressure on the wobbler varies during each cycle, as theindividual pistons change direction. This varying pressure tends tointroduce undesired vibrations into the nutating wobbler motion. Thefull gimbal helps to maintain the nutating-only motion of the wobblerand reduce the undesired vibrations. The velocity of the shoe slippageis restrained by the gimbal-mounted wobbler.

In another embodiment of the present invention, a hold-down plate isused to help maintain the piston shoes against the wobbler of thegimbal-mounted swash-plate. The holes in the hold-down plate areelongated rather than circular. Computer modeling shows that increasingthe elongation of the piston holes in the hold-down plate, especiallythe holes which are the farthest away from the two gimbal anchor points,eliminates impact between the piston shoes and the edges of the holes ofthe hold-down plate.

Although the hydraulic pump and motor of the invention may be used incombination as a stand-alone transmission, the addition of the orbitalgear complex allows a significant decrease in the size of the pump andmotor. In an orbital transmission of the present invention, thehydraulic pump and motor are never doing 100% of the work. The reductionof load on the hydraulics as a result of the gearing also increases thedurability of the pump and motor.

An orbital gear complex is combined with a variable web-rotating deviceto form a minimal orbiter. The variable web-rotating device may be anydevice capable of producing a variable output. An electric generator incombination with an electric motor may be used as the web-rotatingdevice. A variable brake may also be used as the web-rotating device. Ina preferred embodiment, a variable hydraulic pump with a variablehydraulic motor serves as the web-rotating device. The orbital gearingproduces an overdrive condition when the web is stationary. A gearreduction is accomplished by rotating the web with the web-rotatingdevice, which allows for a high gear reduction. In a preferredembodiment, the pump and motor are more preferably long-piston hydraulicmachines with infinitely variable swash plates with full gimbals andelongated holes on the hold-down plate. In a preferred embodiment, apair of hydraulic machines is used in combination with an orbitalgearing system to form a transmission with a minimal orbiter. Thehydraulics effect the gear reduction, and overdrive may be achievedpurely with the orbital gearing.

Long-Piston Hydraulic Machine

Referring to FIG. 1, a variable hydraulic machine 110 includes a modularfixed cylinder block 112. Cylinder block 112 has a plurality ofcylinders 114 (only one shown) in which a respective plurality of matingpistons 116 reciprocate between the retracted position of piston 116 andvariable extended positions (the maximum extension being shown in theposition of piston 116′). Each piston has a spherical head 118 that ismounted on a neck 120 at one end of an elongated axial cylindrical bodyportion 122 that is substantially as long as the length of eachrespective cylinder 114. Each spherical piston head 118 fits within arespective shoe 124 that slides over a flat face 126 formed on thesurface of a rotor 128 that is pivotally attached to a drive element,namely, shaft 130 that is supported on bearings within a bore in thecenter of cylinder block 112.

In one embodiment, hydraulic machine 110 is preferably provided with amodular valve assembly 133 that is bolted as a cap on the left end ofmodular cylinder block 112 and includes a plurality of spool valves 134(only one shown) that regulate the delivery of fluid into and out ofcylinders 114. In another embodiment, a plurality of check valves isalternatively used.

The machine 110 can be operated as either a pump or as a motor. Foroperation as a motor, during the first half of each revolution of driveshaft 130, high pressure fluid from an inlet 136 enters the valve end ofeach respective cylinder 114 through a port 137 to drive each respectivepiston from its retracted position to its fully extended position.During the second half of each revolution, lower pressure fluid iswithdrawn from each respective cylinder through port 137 and fluidoutlet 139 as each piston returns to its fully retracted position.

For operation as a pump, during the one half of each revolution of driveshaft 130, lower pressure fluid is drawn into each respective cylinder114 entering a port 137 from a “closed loop” of circulating hydraulicfluid through inlet 136 as each piston 116 is moved to an extendedposition. During the next half of each revolution, the driving of eachrespective piston 116 back to its fully retracted position directs highpressure fluid from port 137 into the closed hydraulic loop throughoutlet 139. The high pressure fluid is then delivered throughappropriate closed loop piping (not shown) to a mating hydraulicmachine, e.g., hydraulic machine 110 discussed above, causing thepistons of the mating machine to move at a speed that varies with thevolume (gallons per minute) of high pressure fluid being delivered in amanner well known in the art.

The cylindrical wall of each cylinder 114 in modular cylinder block 112is transected radially by a respective lubricating channel 140 formedcircumferentially therein. A plurality of passageways 142 interconnectall lubricating channels 140 to form a continuous lubricating passagewayin cylinder block 112.

Each respective lubricating channel 140 is substantially closed by theaxial cylindrical body 122 of each respective piston 116 during theentire stroke of each piston. That is, the outer circumference of eachcylindrical body 122 acts as a wall that encloses each respectivelubricating channel 140 at all times. Thus, even when pistons 116 arereciprocating through maximum strokes, the continuous lubricatingpassageway interconnecting all lubricating channels 140 remainssubstantially closed off. Continuous lubricating passageway 140, 142 issimply and economically formed within cylinder block 112.

During operation of hydraulic machine 110, all interconnectedlubricating channels 140 are filled almost instantly by a minimal flowof high-pressure fluid from inlet 136 entering each cylinder 114 throughport 137 and being forced between the walls of the cylinders and theouter circumference of each piston 116. Loss of lubricating fluid fromeach lubricating channel 140 is restricted by a surrounding seal 144located near the open end of each cylinder 114. Nonetheless, thelubricating fluid in this closed continuous lubricating passageway oflubricating channels 140 flows moderately but continuously as the resultof a continuous minimal flow of fluid between each of the respectivecylindrical walls of each cylinder and the axial cylindrical body ofeach respective piston in response to piston motion and to the changingpressures in each half-cycle of rotation of drive shaft 130 as thepistons reciprocate. As the pressure in each cylinder 114 is reduced tolow pressure on the return stroke of each piston 116, the higherpressure fluid in otherwise closed lubricating passageway 140, 142 isagain driven between the walls of each cylinder 114 and the outercircumference of body portion 122 of each piston 116 into the valve endof each cylinder 114 experiencing such pressure reduction.

The flow of lubricating fluid in closed continuous lubricatingpassageway 140, 142 is moderate but continuous as the result of asecondary minimal fluid flow in response to piston motion and to thechanging pressures in each half-cycle of rotation of drive shaft 130 asthe pistons reciprocate.

Rotor 128 of pump 110 is pivotally mounted to drive shaft 130 about anaxis 129 that is perpendicular to axis 132. Therefore, while rotor 128rotates with drive shaft 130, its angle of inclination relative to axis130 is preferably variable from 0° (i.e., perpendicular) to +25°. InFIG. 1, rotor 128 is inclined at +25°. This variable inclination iscontrolled as follows: The pivoting of rotor 128 about axis 129 isdetermined by the position of a sliding collar 180 that surrounds driveshaft 130, and is movable axially relative thereto. A control link 182connects collar 180 with rotor 128 so that movement of collar 180axially over the surface of drive shaft 130 causes rotor 128 to pivotabout axis 129. For instance, as collar 180 is moved to the right inFIG. 1, the inclination of rotor 128 varies throughout a continuum fromthe +25° inclination shown, back to 0° (i.e., perpendicular), and thento −25°.

The axial movement of collar 180 is controlled by the fingers 184 of ayoke 186 as yoke 186 is rotated about the axis of a yoke shaft 190 byarticulation of a yoke control arm 188. Yoke 186 is actuated by aconventional linear servo-mechanism (not shown) connected to the bottomof yoke arm 188. While the remaining elements of yoke 186 are allenclosed within a modular swash plate housing 192, and yoke shaft 190 issupported in bearings fixed to housing 192, yoke control arm 188 ispositioned external of housing 192. Swash plate rotor 128 is balanced bya shadow link 194 that is substantially identical to control link 182and is similarly connected to collar 180 but at a location on exactlythe opposite side of collar 180.

Referring to both FIG. 1 and FIG. 2, the cylindrical wall of eachcylinder 114 is transected radially by a respective lubricating channel140 formed circumferentially therein. A plurality of passageways 142interconnect all lubricating channels 140 to form a continuouslubricating passageway in cylinder block 112. Each respectivelubricating channel 140 is substantially closed by the axial cylindricalbody 122 of each respective piston 116 during the entire stroke of eachpiston. That is, the outer circumference of each cylindrical body 122acts as a wall that encloses each respective lubricating channel 140 atall times. Thus, even when pistons 116 are reciprocating through maximumstrokes, the continuous lubricating passageway interconnecting alllubricating channels 140 remains substantially closed off. Continuouslubricating passageway 140, 142 is simply and economically formed withincylinder block 112 as can be best appreciated from the schematicillustration in FIG. 2 in which the relative size of the fluid channelsand connecting passageways and has been exaggerated for clarification.

During operation of hydraulic machine 110, all interconnectedlubricating channels 40 are filled almost instantly by a minimal flow ofhigh-pressure fluid from inlet 36 entering each cylinder 114 throughport 137 and being forced between the walls of the cylinders and theouter circumference of each piston 116. Loss of lubricating fluid fromeach lubricating channel 140 is restricted by a surrounding seal 144located near the open end of each cylinder 114. Nonetheless, thelubricating fluid in this closed continuous lubricating passageway oflubricating channels 140 flows moderately but continuously as the resultof a continuous minimal flow of fluid between each of the respectivecylindrical walls of each cylinder and the axial cylindrical body ofeach respective piston in response to piston motion and to the changingpressures in each half-cycle of rotation of drive shaft 130 as thepistons reciprocate. As the pressure in each cylinder 114 is reduced tolow pressure on the return stroke of each piston 116, the higherpressure fluid in otherwise closed lubricating passageway 140, 142 isagain driven between the walls of each cylinder 114 and the outercircumference of body portion 122 of each piston 116 into the valve endof each cylinder 114 experiencing such pressure reduction.

Referring to FIG. 3A and FIG. 3B, a hold-down assembly for a hydraulicmachine includes a hold-down element 154 with a plurality of circularopenings 160, each of which surrounds the neck 120 of a respectivepiston 116. The swash plate is at +25° angle in FIG. 3A and FIG. 3B.FIG. 3A shows the hold-down plate 154 from the perspective of lookingdown the shaft of the rotor 128, or from plane 3A-3A of FIG. 1. Aplurality of special washers 156 is positioned, respectively, betweenhold-down element 154 and each piston shoe 124. Each washer 156 has anextension 158 that contacts the outer circumference of a respective shoe124 to maintain the shoe in contact with flat face 126 of rotor 128 atall times. Each respective shoe cavity is connected through anappropriate shoe channel 162 and piston channel 164 to assure that fluidpressure present at the shoe-rotor interface is equivalent at all timeswith fluid pressure at the head of each piston 116.

Fluid pressure constantly biases pistons 116 in the direction of rotor128, and the illustrated thrust plate assembly is provided to carry thatload. However, at the speeds of operation required for automotive use(e.g., 4000 rpm) additional bias loading is necessary to assure constantcontact between piston shoes 124 and flat surface 126 of rotor 128. Thevariable hydraulic machines provide such additional bias by using one ofthree simple spring-biased hold-down assemblies.

The first hold-down assembly, for hydraulic machine 110, includes a coilspring 150 that is positioned about shaft 130 and received in anappropriate crevice 152 formed in cylinder block 112 circumferentiallyabout axis 132. Coil spring 150 biases a hold-down element 154 that isalso positioned circumferentially about shaft 130 and axis 132.Hold-down element 154 is provided with a plurality of circular openings160, each of which surrounds the neck 120 of a respective piston 116. Aplurality of special washers 156 is positioned, respectively, betweenhold-down element 154 and each piston shoe 124. Each washer 156 has anextension 158 that contacts the outer circumference of a respective shoe124 to maintain the shoe in contact with flat face 126 of rotor 128 atall times.

The positions of the swash plate and piston shoe hold-down assemblychange relative to each other, as the inclination of rotor 128 isaltered during machine operation. Referring to the relative position ofthese parts at 0° inclination, each piston channel 164 has the sameradial position relative to each respective circular opening 160 inhold-down element 154. At all inclinations other than 0°, the relativeradial position of each piston channel 164 is different for each opening160, and the relative positions of each special washer 156 is alsodifferent. The different relative positions at each of the nine openings160 are themselves constantly-changing as rotor 128 rotates and nutatesthrough one complete revolution at each inclination. For instance, atthe 25° inclination shown in FIG. 3A, if during each revolution of rotor128, one were to watch the movement occurring through only the opening160 at the top (i.e., at 12 o'clock) of hold-down element 154, therelative position of the parts viewed in the top opening 160 wouldserially change to match the relative positions shown in each of theother eight openings 160.

At inclinations other than 0°, during each revolution of rotor 128, eachspecial washer 156 slips over the surface of hold-down element 154 as,simultaneously, each shoe 124 slips over the flat face 126 of rotor 128.Each of these parts changes relative to its own opening 160 through eachof the various positions that can be seen in each of the other eightopenings 160. Each follows a cyclical path (that appears to trace alemniscate, i.e., a “figure-eight”) that varies in size with the angularinclination of swash plate rotor 128 and the horizontal position of eachpiston 116 in fixed cylinder block 112. To assure proper contact betweeneach respective shoe 124 and flat surface 126 of rotor 128, a size ispreferably selected for the boundaries of each opening 160 so that theborders of opening 160 remain in contact with more than one-half of thesurface of each special washer 156 at all times during each revolutionfor all inclinations of rotor 128.

A second hold-down assembly is shown schematically in FIG. 4 in anenlarged, partial, and cross-sectional view of a single piston of ahydraulic machine 210. Each piston 216 is positioned in the modularfixed cylinder block 212 within a cylinder 214, the latter beingtransected radially by a respective lubricating channel 240 formedcircumferentially therein. In the same manner as described in relationto the other hydraulic machines already detailed above, each lubricatingchannel 240 is interconnected with similar channels in the machine'sother cylinders to form a continuous lubricating passageway in cylinderblock 212. An optional surrounding seal 244 may be located near the openend of each cylinder 214 to minimize further the loss of lubricatingfluid from each lubricating channel 240.

Fixed cylinder block 212 includes neither a large axiallycircumferential coil spring nor an axially circumferential crevice forholding same. The modular fixed cylinder block 212 of hydraulic machine210 can be connected to either a modular fixed-angle swash plateassembly or a modular variable-angle swash plate assembly, but in eithercase, hydraulic machine 210 provides a much simpler hold-down assembly.Namely, the hold-down assembly of this embodiment includes only arespective conventional piston shoe 224 for each piston 216 incombination with only a respective coil spring 250, the latter alsobeing associated with each respective piston 216.

Each piston shoe 224 is similar to the conventional shoes shown in thefirst hold-down assembly and is mounted on the spherical head 218 ofpiston 216 to slide over the flat face 226 formed on the surface of themachine's swash plate rotor 228. Each coil spring 250 is, respectively,seated circumferentially about hydraulic valve port 237 at the valve endof each respective cylinder 214 and positioned within the body portionof each respective piston 216.

Each shoe 224 slips over flat face 226 of rotor 228 with a lemniscatemotion that varies in size with the horizontal position of each piston216 and the inclination of rotor 228 relative to axis 232. During normaloperation of hydraulic machine 210, shoes 224 are maintained in contactwith flat face 226 of the swash plate by hydraulic pressure. Therefore,the spring bias provided by coil springs 250 is minimal but sufficientto maintain effective sliding contact between each shoe 224 and flatface 226 in the absence of hydraulic pressure at the valve end of eachrespective cylinder 214. The minimal bias of springs 250 not onlyfacilitates assembly but also prevents entrapment of tiny dirt and metaldetritus encountered during assembly and occasioned by wear.

Referring to FIG. 5, a third hold-down assembly for a hydraulic machine310 includes an improved conventional split swash plate arrangement. Aplurality of pistons 316, each including a respective sliding shoe 324,reciprocates in respective cylinders 314 formed in cylinder block 312that is identical to cylinder block 112. Each shoe 324 slides on theflat face 326 formed on a wobbler 327 that is mounted on a mating rotor328 by appropriate bearings 372, 374 that permit wobbler 327 to nutatewithout rotation while rotor 328 both nutates and rotates in a mannerwell known in the art. The inclination of wobbler 327 and rotor 328about axis 329 is controlled by the position of a sliding collar 380, acontrol link 382, and a balancing shadow link 394.

Shoes 324 are held down by a hold-down assembly substantially identicalto the first hold-down assembly, however, the large single coil spring150 is replaced by a plurality of smaller individual coil springs.

A hold-down plate 354 is fixed to wobbler 327. Each shoe 324 receivesthe circumferential extension of a respective special washer 356, andthe neck of each piston 316 is positioned within one of a correspondingplurality of respective openings 360 formed through hold-down plate 354.While wobbler 327 does not rotate with rotor 328, the nutationalmovement of wobbler 327 is identical to the nutational movement of rotor328 and, therefore, the relative motions between shoes 324 and the flatsurface 326 of wobbler 327 are also identical to those in the firsthold-down assembly.

A plurality of individual coil springs 350 provides the minimal springbias to maintain effective sliding contact between each shoe 324 andflat face 326 of wobbler 327 in the absence of hydraulic pressure at thevalve end of each cylinder 314. Each coil spring 350 is positionedcircumferentially about each shoe 324, being captured between eachspecial washer 356 and a collar formed just above the bottom of eachshoe 324.

Referring to FIG. 6, each hydraulic machine, whether a motor or a pump,is preferably paired with another hydraulic machine, a mating pump ormotor, in a well known “closed loop” arrangement. For example, thehigh-pressure fluid exiting from the outlet 139 of hydraulic machine 110is directly delivered to the input 136′ of a mating hydraulic machine110′, while the low-pressure fluid exiting from the outlet 139′ ofhydraulic machine 110′ is directly delivered to the input 136 of matinghydraulic machine 110. Hydraulic machine 110 and hydraulic machine 110′may be identical in structure except that hydraulic machine 110 is usedas a pump and hydraulic machine 110′ is used as a motor. A portion ofthe fluid in this closed loop system is continually lost to “blow-by”and is collected in a sump, and fluid is automatically delivered fromthe sump back into the closed loop to maintain a predetermined volume offluid in the closed loop system at all times.

Hydraulic Machine with Full Gimbal

During recent developmental work on a long-piston hydraulic machine,vibration at higher speeds and pressures has been noted. Someinterference between the heads of the long pistons and the hold-downplate has also caused the bronze shoes on the pistons to loosen.Repetitive impacts between the bronze shoes and the hold-down plateincrease the operating noise of these hydraulic machines. Although thesehydraulic machines show remarkably low blow-by of lubricating hydraulicfluid in comparison to prior art hydraulic pumps and motors, asignificant amount of this blow-by results from the loosening of thebronze shoes over time. Eliminating the repetitive impacts significantlyimproves the performance of these machines. This undesirableinterference occurs during the relative lemniscate motion shared by thepiston shoes as they slide over the surface of the wobbler portion ofthe split swash plate.

To improve the performance of the hydraulic machines of U.S. Patent App.No. 2004/0168567 the wobbler is further stabilized, and that isaccomplished with a full gimbal. The force applied by the sliding shoeof each piston on the wobbler has both an axial component and a radialcomponent. As the angle of the swash plate increases, the radial forcecomponent increases, and the full gimbal provides structural support tooppose this force and maintain the nutating motion of the wobbler.

A split swash plate of a long piston hydraulic machine includes arotating and nutating “rotor” that is driven by the input axle and anutating-only “wobbler” that rides on the surface of the rotor onbearings. The sliding shoes on the long piston heads move in a “figureeight” path over the surface of the nutating wobbler. However, duringthe relative motion of the sliding shoes on the wobbler, this “figureeight” is really a lemniscate in three dimensions on the surface of animaginary sphere having a diameter that decreases as the angle of theswash increases.

In an embodiment of the present invention, the wobbler is stabilized bya full gimbal. As the wobbler nutates, the distribution of piston shoepressure on the wobbler varies during each cycle, as the individualpistons change direction. This varying pressure tends to introduceundesired vibrations into the nutating wobbler motion. The full gimbalhelps to maintain the nutating-only motion of the wobbler and reduce theundesired vibrations. The velocity of the shoe slippage is directed bythe gimbal-restrained wobbler.

Referring to FIG. 7, a long-piston hydraulic machine 330 with a gimbalis shown. A plurality of pistons 316, each including a respectivesliding shoe, reciprocate in respective cylinders 314 formed in cylinderblock 312. Each shoe 324 slides on the flat face formed on wobbler 327that is mounted on mating rotor 328 by appropriate bearings that permitwobbler 327 to nutate without rotation while rotor 328 both nutates androtates in a manner well known in the art. The inclination of wobbler327 and rotor 328 about axis 329 is controlled by the position ofsliding collar 380, control link 382, and balancing shadow link 394.Shoes are held down by a hold-down assembly substantially identical tothe first hold-down assembly, however, the large single coil spring isreplaced by a plurality of smaller individual coil springs.

The gimbal includes a yoke 332, a first pair of gimbal pins 334connecting yoke 332 to the swash plate housing 392, and a second pair ofgimbal pins 336 connecting yoke 332 to wobbler 327. Yoke 332 forms acomplete annulus around the wobbler. Gimbal pins 334 are located at 180degrees to each other. Gimbal pins 336 are located at 180 degrees toeach other and at 90 degrees to gimbal pins 334. The gimbal structureallows wobbler 327 to nutate but inhibits rotational movement of wobbler327.

A hold-down plate 338 is fixed to wobbler 327. Each shoe receives thecircumferential extension of a respective special washer, and the neckof each piston 316 is positioned within one of a corresponding pluralityof respective openings formed through hold-down plate 338. While wobbler327 does not rotate with rotor 328, the nutational movement of wobbler327 is identical to the nutational movement of rotor 328 and, therefore,the relative motions between shoes and the flat surface of wobbler 327are also identical to those in the first hold-down assembly.

A plurality of individual coil springs provides the minimal spring biasto maintain effective sliding contact between each shoe and the flatface of wobbler 327 in the absence of hydraulic pressure at the valveend of each cylinder 314. Each coil spring is positionedcircumferentially about each shoe, being captured between each specialwasher and a collar formed just above the bottom of each shoe.

In another embodiment of the present invention, the holes in thehold-down plate are elongated rather than circular. The full gimbal isanchored at two points around the wobbler. Computer modeling shows thatincreasing the elongation of the piston holes in the hold-down plate,especially the holes which are the farthest away from the two anchorpoints, eliminates impact between the piston shoes and the edges of theholes of the hold-down plate.

Referring to FIG. 8, hold-down plate 338 preferably has one opening 340directly aligned with one of two gimbal pins 336 connecting yoke 332 towobbler 327. The relative locations of gimbal pins 334, 336 in relationto hold-down plate 338 are shown schematically in FIG. 8. Opening 340 isnearly circular. The shapes of the openings 340, 342, 344, 346, 348 aremore elongated the farther the openings are from gimbal pins 336.

Computer models of this new design indicate it is possible to use only acombination of two hydraulic machines, one as a pump and one as a motor,in a closed hydraulic loop as an infinitely variable transmission formany models of present day vehicles. However, the inventors believe thatthe addition of the described orbital gear complex provides a moreefficient transmission than these hydraulic machines acting alone. Thishydraulic/orbital gear transmission provides a significantly improvedautomotive transmission that can be scaled up or down to meet a widespectrum of weight and size requirements.

Orbital Transmission

Although the hydraulic pump and motor may be used in combination as astand-alone transmission, the addition of the orbiter allows asignificant decrease in the size of the pump and motor. In an orbitaltransmission of the present invention, the hydraulic pump and motor arenever doing 100% of the work. The reduction of load on the hydraulics asa result of the gearing also increases the durability of the pump andmotor.

A transmission 400 is shown in FIG. 9 in an embodiment of the presentinvention. An engine 402 is shown connected to the invention'stransmission that includes only a minimal orbiter 404 and a variableweb-rotating device 403. In one embodiment, variable web-rotating device403 includes an electric motor 406 in combination with an electricgenerator 408 with an electrical connection 440 between them. In anotherembodiment, variable web-rotating device 403 includes a variablehydraulic motor 406 in combination with a variable hydraulic pump 408with a hydraulic connection 440 between them. The transmission isdescribed below with a hydraulic pump and a hydraulic motor as apreferred web-rotating device.

Orbiter 404 includes only an input gear 410 and an output gear 412, bothmounted for rotation about a first axis 414, and a cluster gear 416mounted for rotation about a second axis 418 parallel to first axis 414.Input gear 410 is fixed for rotation with the drive shaft 420 of engine402, while output gear 412 is fixed for rotation with an output shaft422. Cluster gear 416 is fixed to an orbit shaft 424 supported forrotation in a web 426, and web 426 is itself mounted to rotate aboutfirst axis 414, thereby permitting orbit shaft 424 and cluster gear 416to orbit, respectively, about first axis 414, as well as about inputgear 410 and output gear 412. Cluster gear 416 has two sets of gearteeth 428, 430 that mesh, respectively, with the teeth of input gear 410and output gear 412.

The gear tooth ratios between input gear 410 and cluster gear 428, andbetween cluster gear 430 and output gear 412, are selected so that, whenrotation of web 426 is prevented, output gear 412 rotates at apredetermined overdrive of rotation of input gear 410. This is incontrast to the teachings of the prior art, where a gear reduction istaught, as in the orbital gearing of previously cited U.S. Pat. No.6,748,817. For instance, in a preferred embodiment of the presentinvention, gear tooth ratios are selected as shown in Table 1.

TABLE 1 Gear No. of Teeth Input gear 410 36 Cluster gear 428 27 Clustergear 430 36 Output gear 412 27

With this gearing example, when rotation of web 426 is prevented, outputgear 412 rotates at an overdrive of approximately 0.6:1 of the rotationof input gear 410.

Fixed to the outside of web 426 is a gear 432 that meshes with a motorgear 434 that is connected to the motor shaft 436. Motor shaft 436 maybe made disconnectable to motor gear 434 by an optional first clutch 438that, preferably, is a simple jaw clutch, but may be any type of clutch.Clutch 438 provides a true-neutral safety feature by assuring that thevehicle is in neutral, especially important at start-up, regardless ofthe output of variable web-rotating device 403. For the purpose ofillustration, motor shaft 436 is driven by hydraulic control motor 406and rotates motor gear 434 and web gear 432 in a 1:1 relationship. Theratio between the motor gear 434 and the web gear 432 may fall within arange without deviating from the spirit of the invention. This ratio,along with the gear tooth ratio, may be varied to produce particularinput-to-output transmission gear ratios for particular settings ofvariable web-rotating device 403. Control motor 406 is operated, inturn, by hydraulic fluid delivered from hydraulic pump 408 through a“closed-loop” hydraulic circuit 440. An auxiliary drive gear 442 that isfixed to engine drive shaft 420 causes the rotation of a first matinggear 444 and a pump shaft 446 in a 1:1 relationship.

Although the operation of hydraulic pump/motor combinations is wellknown in the art, operation of a pump/motor especially suited for thistransmission is discussed in detail in this disclosure. The auxiliaryrotation of pump shaft 446 by engine drive shaft 420 permits hydraulicpump 408 to create a flow of hydraulic fluid for control motor 406 inaccordance with the adjusted angle of the swash plate (not shown) ofpump 408.

Forward Operation of IVT

As an illustration of the operation of a transmission of the presentinvention, the orbital gear tooth ratios set forth above are used hereto calculate the values given Table 2. Table 2 shows the conditions ofthe pump and motor swash plates, the resulting web rotation rate, outputshaft rotation rate and transmission ratio at discrete stages fromreverse to neutral to overdrive. However, it should be understood thatthe infinitely variable transmission goes through a continuum oftransmission ratios over this entire range. An idling speed of 500 RPMis used for the sample calculations.

TABLE 2 Vehicle Engine Pump Motor Web Output Transmission Speed RPMSwash Swash RPM RPM Ratio Reverse 500 25° 10° 1250 −83 −6:1  Neutral 50025° 10.9°   1143 - 0 - 1:0 Low 1 500 25° 13.5°   929 167 3:1 Low 2 50025° 15.2°   821 250 2:1 Drive 500 25° 25° 500 500 1:1 Overdrive 1 50012.5°   25° 250 694 0.7:1   Overdrive 2 500  0° 25° - 0 - 889 0.6:1  Overdrive 3 500 −7.1°   25° −143 1000 0.5:1  

With reference to Table 2, is should be initially noted that as theangle of the swash plate of motor 406 is continuously increased in apositive direction, the rate of rotation of motor shaft 436 graduallyslows, thereby slowing the rotation of web 426 and causing output shaft422 to slowly increase its rate of rotation.

With control motor 406 connected to web 426 and the swash plate of pump408 set at its maximum inclination (i.e. 25°), setting the swash plateangle of motor 406 to 10.9° causes web 426 to rotated at a speed thatcauses output shaft 422 to come to a stop, i.e., to what is, in effect,a “geared neutral” condition.

To change from “Neutral” condition to “Drive”, the swash angle of pump408 is held at 25° while the swash angle of motor 406 is continuouslyincreased to 25°, where the transmission ratio equals 1:1. From “Drive”to a predetermined “Overdrive 2”, the swash angle of motor 406 swash isheld at 25° while the swash angle of pump 408 is continuously decreasedto 0°. When the swash plate of pump 408 reaches 0°, web 426 is stopped,and the speed of rotation of output gear 412 and output drive shaft 422is greater than the speed of engine drive shaft 420 and input gear 410by the overdrive predetermined by the basic gear complex referred toabove.

As an important additional feature of the present invention, the swashangle of pump 408 may be decreased to a slightly negative angle, therebyreversing the direction of motor shaft 436 to extend the invention's“infinite overdrive” throughout a range beyond “Overdrive 2”.

Therefore, as the speed of rotation of motor 406 and web 426continuously decreases, the forward rotation of output shaft 422continuously increases in speed through an infinite range of gear ratiosfrom a high gear reduction (significantly greater than 3:1) to 1:1 andthen on through an extended continuous overdrive without any gearshifting, clutch shifting, or any significant fluctuation in enginespeed.

During any required speed reduction, such as braking, the swash plateangles are adjusted in the opposite direction toward the geared neutralposition to achieve an appropriate gear reduction. Similarly, ifadditional power is required when the vehicle is in full overdrive, suchas for climbing a hill or passing, the pump swash plate angle may beincreased to provide an appropriate gear reduction or, of course, theengine speed may be increased.

Special attention is called to the fact that this just-describedcontinuous and infinite-progression gear ratio change (from start-up tooverdrive) occurs without any significant change in the speed of engine402. The engine may be maintained at a relatively low and efficientoperational level throughout the entire acceleration from a standingstop up to overdrive. This remarkable feature not only results in fuelsavings but, more importantly, in significant reduction in pollution.This is particularly true for diesel engine vehicles, since the engine'sselected operational speed can be predetermined at a “sweet spot” whichoptimizes performance. As is well known, when a diesel engine operatesat a constant speed, it discharges little, if any, pollutants.

A further feature of the invention provides a tow/haul mode to achievehigher fuel efficiency and to reduce the duty cycle on the hydraulicswhen the vehicle is hauling heavy loads, towing a trailer, or travellingover steep terrain at highway speeds. This feature by-passes thehydraulic system and locks up the drive in a geared 1:1 with the vehicleengine. This is accomplished by the addition of a second clutch 439 thatis used to engage a further gear 435 to pump shaft 446. Gear 435 isinterconnected in a gear train with, and is the same size as, auxiliarydrive gear 442, first mating gear 444, and web gear 432. Therefore, whenfirst clutch 438 disengages motor 406 from motor gear 434 and secondclutch 439 engages gear 435 to pump shaft 446, the direct gear trainfrom engine 402 through gears 442, 444, 435 and 432, rotates web 426 atthe same speed that engine 402 is rotating input gear 410. This resultsin the direct geared drive of both output gear 412 and output driveshaft 422 in a 1:1 relationship with engine 402.

“Stopping” and Rearward Operation of IVT

When control motor 406 is connected to web 426, and the swash plate ofpump 408 is set at its maximum inclination (i.e. 25°), the swash plateof motor 406 is set to achieve “Neutral” (i.e., at 10.90 for thepreviously-given gear ratios). Under these conditions, output shaft 422is stationary. As indicated above, this in effect provides a “gearedneutral” in which web 426 is held by a constant torque in a stoppedposition for start up and when starting in traffic. However, it shouldbe noted that at any time under these conditions, clutch 438 can bedisengaged and a “true neutral” can be achieved to disconnect the drivefrom the vehicle wheels completely.

If the angle of the swash plate of motor 406 is moved in a slightlynegative direction from its geared neutral setting (e.g., by 1-3°),control motor 406 continues to rotate web 426 in the same direction thatachieves “geared neutral”, but web 426 rotates at a slightly fasterspeed. The net effect is that output shaft 422 now rotates at arelatively high gear reduction in the rearward direction, i.e., in“Reverse”.

When the setting of the swash plate of motor 406 is continuouslyincreased in a negative direction (i.e., beyond the setting used tobring output gear 412 to a stop), rotations of web 426, output gear 412,and output shaft 422 all continuously increase in the forward direction.If control motor 406 is “neutralized” (e.g., by disengaging clutch 438),an idling-speed rotation of input gear 410 will automatically causecluster gear 416 to rotate web 426 in the rearward direction at theexact speed that causes output gear 412 to come to a complete stop. Thatis, when rotational control of the web is neutralized, the minimalorbiter of this invention automatically seeks the position of minimumtorque.

Therefore, it may not be necessary to program precisely the adjustmentof the swash plate of motor 406 in order create the requiredpredetermined speed reversal of the web for bringing the transmission tozero speed when stopping the vehicle. We have developed a preferredhydraulic pump/motor embodiment for the invention that, without a firstclutch 438, still permits control motor 406 to be adjusted appropriatelyto allow the vehicle to come to a complete stop whenever the speed ofinput gear 410 is reduced to idling engine speed.

Power Takeoff

As is well known in the art, power takeoff shafts are often provided ontractors and trucks to permit auxiliary equipment to be operated fromthe vehicle's engine. Therefore, one other feature of the transmissionis a power takeoff assembly 450 that includes a power takeoff shaft 452and a power takeoff gear 454 connected by a clutch 456.

Power takeoff gear 454 is driven by auxiliary drive gear 442. Powertakeoff gear 454 generally “free-wheels”, being disconnected from powertakeoff shaft 452 by normally disengaged clutch 456. However, whenclutch 456 is engaged, power takeoff shaft 452 also rotates to operateauxiliary equipment.

Hydraulic By-Pass Circuit

A valve-regulated “by-pass” assembly is preferably incorporated in theclosed-loop hydraulic circuitry 440 shared by the hydraulic pump 408 andmotor 406. (Such a by-pass arrangement is disclosed in above-referencedU.S. Pat. No. 6,748,817. A pair of “by-pass” passageways connects theopposite sides of the closed-loop and pass through a cylinder, beingblocked by the piston portions of a spool valve. A pair of stems islocated on the spool valve so that, when the spool valve is moved in onedirection, the stems permit hydraulic fluid to flow through by-passpassageways. A sensor is responsive to upper and lower levels inselected parameters of vehicle operation (e.g., vehicle speed and/orhydraulic pressure in the closed-loop). Sensing a first level of theseselected parameters causes the spool valve to move in one direction toopen the passageways (e.g., whenever the vehicle speed is reduced andapproaches a stopped condition), while sensing a second level restoresthe valve to the opposite position, returning the closed-loop hydrauliccircuit 440 to its normal condition.

Activation of the spool valve to open the by-pass circuit permits shaft436 of control motor 406 to be moved independently even though the swashplate of pump 408 is being driven or is being held stopped at 0°.Therefore, the by-pass assembly can be used to reduce the transmissionload during engine start-up, thereby replacing a vehicle's fly-wheelclutch. In this regard, since the sensor can be used to sense asignificant change in fluid pressure in the closed-loop hydrauliccircuit 440, the by-pass assembly can also serve as a safety device,preventing any exceptional overload of the hydraulic system.

As a further feature of an orbital transmission of the presentinvention, the efficiency of the hydraulic pump and motor in combinationwith the orbital gearing allows sufficient power to be transmitted tothe wheels, while the engine speed is maintained around 500 RPM. Thisfeature provides significant fuel savings despite the fact that theengine runs well below the “sweet spot” for which transmissions of theprior art are designed. The sweet spot of an engine is the region of theefficiency map where the engine is most efficient at converting fuelinto mechanical power. For most automobile engines, the sweet spot isfound in the range of 1500 RPM. A transmission of the present inventionimproves fuel economy despite maintaining the engine at a speed withreduced efficiency. The loss in efficiency is more than compensated bythe reduced fuel requirements of running at 500 RPM. A further advantageof running at such low engine speeds is a reduced pressure demand on thehydraulic pump and motor, thereby reducing the duty cycle on thehydraulics and further improving their durability.

In terms of fuel economy, an orbital transmission of the presentinvention brings city driving efficiency up to highway drivingefficiency. Since city driving accounts for about 60% of all driving,incorporation of a transmission of the present invention intoautomobiles would provide a significant fuel savings. Today's enginesare designed to run most efficiently in the range of 1500 RPM. Ofcourse, further fuel savings may be achieved by combining a transmissionof the present invention with an automobile engine designed to run mostefficiently around 500 RPM.

A transmission of the present invention is capable of varying the speedof the drive shaft with minimal changes to engine speed. Thus, thepresent invention allows engine speed to remain in a relatively narrowlow-to-moderate range where the combustion in HCCI engines is moreeasily controlled. In a preferred embodiment, the engine 402 is an HCCIengine. A transmission of the present invention is highly compatiblewith implementation of more fuel efficient HCCI engines ongasoline-powered vehicles.

A final feature of the invention is now described with reference to FIG.10A and FIG. 10B, which schematically illustrate the invention's modularhydro-mechanical transmission in place behind a standard automotiveengine 402. In the same general arrangement just described above, therotation of output shaft 414 of the invention's single orbiter 404 isregulated by the hydraulics of motor 406 and pump 408 which, in turn,are driven by engine 402. As shown, orbiter 404 is modularly mounted onplate 404 a which is bolted to the fly-wheel housing 502 at the rear ofengine 402. Similarly, motor 406 and pump 408 are modularly mounted onplate 408 a.

For some vehicles, the modular combination of motor 406 and pump 408 canfulfill the full drive requirements of the vehicle. In suchcircumstances, modular orbiter 404 can be omitted, and, with minormodifications, motor 406, pump 408, and plate 408 a can be modularlybolted directly to the fly-wheel housing 502 at the rear of engine 402.

Accordingly, it is to be understood that the embodiments of theinvention herein described are merely illustrative of the application ofthe principles of the invention. Reference herein to details of theillustrated embodiments is not intended to limit the scope of theclaims, which themselves recite those features regarded as essential tothe invention.

1-12. (canceled)
 13. A method of infinitely varying the speed and torqueof a transmission of a vehicle from a drive shaft of a primary engine toan output shaft comprising the steps of: a) providing gearing to producea predetermined overdrive of the output shaft relative to the driveshaft; b) rotating an orbiter web in a first direction at apredetermined speed to produce no rotation of the output shaft; and c)increasing the speed of the output shaft relative to the drive shaft bydecreasing the speed of rotation of the web.
 14. The method of claim 13further comprising the step of rotating the web in a second directionopposite the first direction to increase the speed of rotation of theoutput shaft relative to the input shaft beyond the predeterminedoverdrive.
 15. The method of claim 13 further comprising the step ofrotating the web faster than the predetermined speed to reverse thedirection of rotation of the output shaft.
 16. A hydromechanicaltransmission for a primary engine, the transmission comprising: avariable hydraulic component driven by a drive shaft of said engine andproducing a hydraulic output; and a mechanical component driven by saiddrive shaft and producing a mechanical output; wherein said hydraulicoutput and said mechanical output are combined to drive an output shaftfor driving a vehicle such that increasing the speed of the hydrauliccomponent decreases the ratio of the speed of said output shaft withrespect to the speed of said drive shaft.
 17. The transmission of claim16, wherein said transmission provides power to said output shaft withthe engine running at a speed well below a conventional optimalefficiency engine speed.
 18. A hydromechanical transmission for aprimary engine having a primary drive shaft and an auxiliary driveshaft, the transmission comprising: a variable hydraulic componentdriven by the auxiliary drive shaft and producing a hydraulic output;and a mechanical component driven by the primary drive shaft andproducing a mechanical output; wherein the hydraulic output and themechanical output are combined to drive an output shaft for driving avehicle; wherein the transmission provides power to the drive shaft withthe engine running at a speed below a conventional optimal efficiencyengine speed. 19-35. (canceled)
 36. The transmission of claim 16,wherein said transmission provides power to said output shaft with theengine running at an idling speed.
 37. The transmission of claim 16,wherein said variable hydraulic component comprises a variable hydraulicmotor driven by a variable hydraulic pump that is driven by said engine.38. The transmission of claim 16, wherein said mechanical componentcomprises an orbiter comprising: an input gear mounted on a first axisand responsive to an input drive provided by the primary engine; anoutput gear mounted on said first axis; and at least one cluster gearmeshed with only said input and output gears and mounted for rotation onan orbit shaft positioned parallel with said first axis.
 39. Thetransmission of claim 38, wherein said mechanical component furthercomprises an orbiter web supporting said orbit shaft and mounted forrotation about said first axis to permit the orbit shaft and the clustergear to orbit, respectively, said first axis and said input and outputgears.
 40. The transmission of claim 39, wherein the gear tooth ratiosbetween said cluster gear and said input and output gears are selectedsuch that, when rotation of the web is prevented, rotation of said inputgear produces rotation of said output gear at a predetermined overdriveof the input drive.
 41. The transmission of claim 18, wherein saidtransmission provides power to said output shaft with the engine runningat an idling speed.
 42. The transmission of claim 18, wherein saidvariable hydraulic component comprises a variable hydraulic motor drivenby a variable hydraulic pump that is driven by said engine.
 43. Thetransmission of claim 18, wherein said mechanical component comprises anorbiter comprising: an input gear mounted on a first axis and responsiveto an input drive provided by the primary engine; an output gear mountedon said first axis; and at least one cluster gear meshed with only saidinput and output gears and mounted for rotation on an orbit shaftpositioned parallel with said first axis.
 44. The transmission of claim43, wherein said mechanical component further comprises an orbiter websupporting said orbit shaft and mounted for rotation about said firstaxis to permit the orbit shaft and the cluster gear to orbit,respectively, said first axis and said input and output gears.
 45. Thetransmission of claim 44, wherein the gear tooth ratios between saidcluster gear and said input and output gears are selected such that,when rotation of the web is prevented, rotation of said input gearproduces rotation of said output gear at a predetermined overdrive ofthe input drive.